This application is based on Application No. 2001-137642, filed in Japan on May 8, 2001, the contents of which are hereby incorporated by reference.
1. Field of the Invention
The present invention relates in general terms to a valve timing control system for an internal combustion engine for controlling operation timings of intake valves and exhaust valves of the engine in dependence on engine operating states.
2. Description of Related Art
In recent years, the statutory regulations imposed in connection with emission of harmful materials or substances contained in the exhaust gas discharged to the atmosphere from the internal combustion engine mounted on a motor vehicle or automobile become more and more severe from the standpoint of environmental protection. Under the circumstances, there exists a great demand for reducing the emission of harmful materials or substances contained in the exhaust gas of the internal combustion engine.
In general, there have heretofore been known two sorts of methods of reducing the harmful exhaust gas components. One method is directed to reduction of the harmful gas directly discharged from the internal combustion engine (hereinafter also referred to simply as the engine) while the other method is directed to the reduction of the harmful components through posttreatment of the engine exhaust gas with the aid of a catalytic converter (hereinafter also referred to simply as the catalyst) installed within the exhaust pipe of the engine at an intermediate portion.
As is well known in the art, in the catalyst such as mentioned above, reaction of rendering the harmful gas components to be harmless is difficult or unable to take place unless the temperature of the catalyst has reached a predetermined value. Consequently, it is an important requirement to increase or rise speedily the temperature of the catalyst even when the internal combustion engine is, for example, in the course of starting operation in the cold state (i.e., in the state of low temperature).
In this conjunction, it is also known that in most of the internal combustion engines known heretofore, cam shafts which plays an essential role in determining the timings for opening and closing the intake or exhaust valves are so arranged as to be rotationally driven by a crank shaft through the medium of timing belts (or timing chains).
Accordingly, the timings for opening and closing the intake or exhaust valves (which timing may also be referred to as the cam angles) are so controlled as to remain constant relative to the crank angle notwithstanding of the fact that the valve timings as required may change in dependence on the operating states of the engine.
However, in recent years, a valve timing control system designed to be capable of changing or modifying the valve timings has been adopted for practical applications with a view to enhancing the fuel-cost performance of the engine while ensuring improvement of the exhaust gas quality.
The valve timing control system of this type is disclosed in, for example, in Japanese Patent Application Laid-Open Publication No. 324613/1997 (JP-A-9-324613).
The valve timing control system disclosed in the above-mentioned publication includes a variable valve timing mechanism (also referred to as the WT mechanism in short) which is comprised of vanes each disposed rotatably within a housing for changing the phase (or angular position) of the cam shafts which is adapted to drive the intake valves and the exhaust valves. Incidentally, concerning arrangement of the vanes, description will be made later on.
At this juncture, however, it should be mentioned that in the engine starting operation, the vane of the variable valve timing mechanism is held substantially at a mid position (start corresponding position) for controlling or regulating the relative angular displacement of the cam angle relative to the crank angle and releasing the regulation or control after lapse of a predetermined time.
For having better understanding of the concept underlying the present invention, description will first be made in some detail of a hitherto known or conventional valve timing control system of an internal combustion engine. FIG. 11 is a functional block diagram showing generally and schematically a configuration of a conventional valve timing control system of an internal combustion engine together with several peripheral parts of the engine.
Referring to FIG. 11, provided in association with an intake pipe 4 for feeding the air into a combustion chamber(s) defined within the cylinder(s) of the engine 1 are an air cleaner 2 for purifying the intake air, an air flow sensor 3 for measuring the quantity or flow rate of the intake air. Further, installed in the intake pipe 4 are a throttle valve 5 for adjusting or regulating the intake air quantity (i.e., the amount or flow rate of the intake air) to thereby control the output of the engine 1, an idle speed control valve (also referred to simply as the ISCV in short) 6 for adjusting or regulating the intake air flow which bypasses the throttle valve 5 to thereby effectuate the engine rotation speed (rpm) control in the idling operation mode, and a fuel injector 7 for charging or injecting an amount of fuel which conforms with the intake air quantity.
Additionally, provided internally of the combustion chamber of the engine cylinder 1 is a spark plug 8 for producing a spark discharge for triggering combustion of the air-fuel mixture charged in the combustion chamber defined within the cylinder. To this end, the spark plug 8 is electrically connected to an ignition coil 9 which supplies electric energy of high voltage to the spark plug 8.
An exhaust pipe 10 is provided for discharging an exhaust gas resulting from the combustion of the air-fuel mixture within the engine cylinder. An O2-sensor 11 and a catalytic converter 12 are disposed in the exhaust pipe 10. The O2-sensor 11 serves for detecting the content of residual oxygen contained in the exhaust gas.
The catalytic converter or catalyst 12 is constituted by a three-way catalytic converter known by itself is capable of eliminating simultaneously harmful gas components such as HC (hydrocarbon), CO (carbon monoxide) and NOx (nitrogen oxides) contained in the exhaust gas.
A sensor plate 13 designed for detecting the crank angle is mounted on a crank shaft (not shown) so as to corotate therewith. The sensor plate 13 is provided with a projection (not shown) at a predetermined crank angle in the outer periphery thereof.
A crank angle sensor 14 is installed at a position diametrically opposite to the outer periphery of the sensor plate 13 for the purpose of detecting the angular position of the crank shaft in cooperation with the sensor plate 13. Thus, the crank angle sensor 14 can generate an electric signal indicative of the crank angle, i.e., the crank angle signal, every time the projection of the sensor plate 13 passes by the crank angle sensor 14. In this way, the rotating position or angular position (crank angle) of the crank shaft can be detected.
The engine 1 is equipped with valves for putting into communication the intake pipe 4 and the exhaust pipe 10 to each other, wherein the timings for driving the intake or exhaust valves are determined by the cam shafts which are rotated at a speed equal to a half of that of the crank shaft, as will be described later on.
Actuators 15 and 16 for changing adjustably the cam phases are designed to change the timings for driving or actuating the intake or exhaust valves, respectively.
More specifically, each of the actuators 15 and 16 is comprised of a retarding hydraulic chamber and an advancing hydraulic chamber partitioned from each other (described later on) for changing or varying the rotational or angular positions (phases) of the cam shafts 15C and 16C, respectively, relative to the crank shaft.
Cam angle sensors 17 and 18 are disposed at positions diametrically opposite to the outer periphery of cam angle detecting sensor plates (not shown) for the purpose of detecting the angular positions of the cams(i.e., cam angles or phases) through cooperation with the sensor plate. More specifically, each of the cam angle sensors 17 and 18 is designed to generate a pulse signal indicative of the cam angle (i.e., the cam angle signal) in response to a projection formed in the outer periphery of the associated cam angle detecting sensor plate in a similar manner as the crank angle sensor 14 described previously. In this way, it is possible to detect the cam angles (or angular position of the cam shafts).
Oil control valves (also referred to as the OCV in short) 19 and 20 constitute hydraulic pressure supply units in cooperation with oil pumps (not shown) and serve for controlling or regulating the hydraulic pressure supplied to the individual actuators 15 and 16 for thereby controlling the cam phases. Parenthetically, the oil pump is designed to feed oil at a predetermined hydraulic pressure.
An electronic control unit (also referred to simply as the ECU) 21 which may be constituted by a microcomputer or microprocessor serves as a control means for the internal combustion engine system. Among others, the ECU 21 is in charge of controlling the fuel injectors 7 and the spark plugs 8 as well as the cam phases (angular positions of the cams) of the actuators 15 and 16 in dependence on the engine operating states detected by the various sensors such as the air-flow sensor 3, the O2-sensor 11, the crank angle sensor 14 and the cam angle sensors 17 and 18.
Further, provided in association with the throttle valve 5 is a throttle position sensor (not shown in the figure) for detecting the throttle opening degree while a water temperature sensor is provided for the engine 1 for detecting the temperature of cooling water therefor. The throttle opening degree and the cooling water temperature as detected are also inputted to the ECU 21 as the information indicative of the operating state of the engine 1 similarly to the various sensor information mentioned above.
Next, description will be made of the conventional engine control operation performed by the prior art valve timing control system shown in FIG. 11. Firstly, the air flow sensor 3 measures the air quantity (flow rate of the intake air) fed to the engine 1, the output of the air-flow sensor 3 being supplied to the ECU 21 as the detection information indicative of the operating state of the engine.
The electronic control unit or ECU 21 arithmetically determines the fuel quantity or amount which conforms to the air quantity as measured to thereby drive or actuate correspondingly the fuel injector 7. At the same time, the ECU 21 controls the time duration for electrical energization of the ignition coil 18 as well as the timing for interruption thereof to thereby produce a spark discharge at the spark plug 8 for igniting or firing the air-fuel mixture charged within the combustion chamber defined within the engine cylinder at a proper timing.
On the other hand, the throttle valve 5 serves for adjusting or regulating the amount of intake air fed to the engine to thereby control correspondingly the output torque or power generated by the engine 1. The exhaust gas resulting from the combustion of the air-fuel mixture within the cylinder of the engine 1 is discharged through the exhaust pipe 10.
In that case, the catalytic converter 12 disposed within the exhaust pipe 10 at an intermediate location thereof converts the harmful components contained in the exhaust gas such as hydrocarbon (HC) (unburned gas), carbon monoxide (CO) and nitrogen oxides (NOx) into harmless carbon dioxide and water (H2O). In this way, the engine exhaust gas is purified.
In order to make available the maximum purification efficiency of the three-way catalytic converter 12, the O2-sensor 11 is installed in association with the exhaust pipe 10 for detecting the amount of residual oxygen contained in the exhaust gas. The output signal of the O2-sensor 11 is inputted to the electronic control unit or ECU 21 which responds thereto by regulating in a feedback loop the amount of fuel injected through the fuel injector 7 so that the air-fuel mixture which is to undergo the combustion can assume the stoichiometric ratio.
In addition, the ECU 21 controls the actuators 15 and 16 (which constitute parts of the variable valve timing mechanism) in dependence on the engine operating state for regulating the timings at which the intake or exhaust valves are to be driven or actuated.
In the following, referring to FIGS. 12 to 13, description will be made in concrete of the phase angle control operation preformed for the cam shafts 15C and 16C by the conventional valve timing control system for the internal combustion engine.
By the way, in the case of the conventional internal combustion engine of the fixed valve timing scheme (not shown), torque of the crank shaft is transmitted to the cam shafts through the medium of the timing belts (timing chains) and transmission mechanisms including pulleys and sprockets and coupled operatively to the cam shafts for corotation with the pulleys.
By contrast, in the case of the internal combustion engine equipped with the variable valve timing mechanism, there are provided the actuators which are designed to change the relative phase position between the crank shaft and the cam shafts in place of the pulleys and the sprockets mentioned above.
FIG. 12 is a view for illustrating relation between the crank angle [xc2x0 CA] and the valve lift stroke (indicating the degree of valve opening [mm], (hereinafter also referred to as the valve opening quantity). In the figure, the top dead center in the compression stroke of the cylinder is designated by reference symbol TDC.
In FIG. 12, a single-dotted broken line curve represents change of the valve lift stroke delimited mechanically in the most retarded state, a broken line curve represents change of the valve lift stroke delimited mechanically in the most advanced state, and a solid line curve represents change of the valve lift stroke in a locked state set by a locking mechanism (described hereinafter).
Referring to FIG. 12, it is to be noted that the peak position of the valve lift stroke on the retarded side (right-hand side as viewed in the figure) with reference to the top dead center (TDC) corresponds to the fully opened position of the intake valve while the peak position of the valve lift stroke on the advanced side (left-hand side as viewed in the figure) corresponds to the fully opened position of the exhaust valve.
Accordingly, difference in the crank angle between the peaks on the retarded side and the advanced side (i.e., difference between the single-dotted line curve and the broken line curve) represents the range within which the valve timing can be changed (i.e., valve timing adjustable range). To say in another way, the valve timing can be changed or adjusted within the crank angle range defined between the broken line curve and the single-dotted line curve in either of the suction and exhaust operation.
FIG. 13 is a timing chart for illustrating phase or timing relations between the output pulse signal of the crank angle sensor 14 on one hand and that of the cam angle sensor 17 or 18 on the other hand. More specifically, shown in FIG. 13 are the output pulse signals of the cam angle sensor 17 or 18 in both the most retarded state and the most advanced state, respectively, relative to the output of the crank angle sensor.
In this conjunction, it should be added that the phase position of the output signal of the cam angle sensor 17 or 18 relative to the output signal of the crank angle sensor 14 (i.e., crank angle signal) becomes different in dependence on the positions at which the cam angle sensors 17 and 18 are mounted.
At this juncture, it should further be mentioned that retarding of the valve timing means that the valve opening start timings of both the intake or exhaust valves is retarded or delayed relative to (or with reference to) the crank angle, while advancing of the valve timing means that the valve opening start timings of both the valves is advanced relative to the crank angle.
The opening start timings for the intake valve and the exhaust valves can be changed or modified by means of the actuators 15 and 16 which constitute parts of the variable valve timing mechanism to be thereby so controlled as to assume a given retarded position or advanced position within the aforementioned valve timing adjustable or variable range mentioned hereinbefore by reference to FIG. 12.
FIGS. 14 to 16 are views showing internal structures of the actuators 15 and 16 which are implemented in a substantially identical structure. More specifically, FIG. 14 shows the same in a state where the cam phase is adjusted to the most retarded position (corresponding to the state indicated by the single-dotted line curve in FIG. 12), FIG. 15 shows the same in a state where the cam phase is adjusted to the locked or lock-up position (corresponding to the state indicated by the solid line curve in FIG. 12), and FIG. 16 shows the same in a state where the cam phase is adjusted to the most advanced position (corresponding to the state indicated by the broken line curve in FIG. 12), respectively.
Referring to FIGS. 14, 15 and 16, each of the actuators 15 and 16 is comprised of a housing 151 which is rotatable in the direction indicated by an arrow, a vane 152 rotatable together with the housing 151, retarding hydraulic chambers 153 and advancing hydraulic chambers 154 both defined internally of the housing 151, a lock pin 155 and a spring 156 which are also provided within the housing 151, and locking recesses 157 formed in the vane 152.
Power or torque is transmitted to the housing 151 from the crank shaft through the medium of a belt/pulley assembly (not shown) with the speed of rotation being reduced by a factor of 1/2.
The position (phase position) of the vane 152 is caused to shift within the housing 151 in response to the hydraulic pressure supplied selectively to the retarding hydraulic chamber 153 or the advancing hydraulic chamber 154.
The range of operation (hereinafter also referred to as the operation range) of the vane 152 is determined or defined by the retarding hydraulic chamber 153 and the advancing hydraulic chamber 154.
The spring 156 resiliently urges the lock pin 155 in the protruding direction while the locking recess 157 is formed at a predetermined vane lock-up position so that the recess 157 faces in opposition to the tip end of the lock pin 155.
Parenthetically, an oil feed port (not shown) is formed in the locking recess 157 through which the hydraulic medium (i.e., oil in this case) is supplied interchangeably from either one of the retarding hydraulic chamber 153 and the advancing hydraulic chamber 154 within which a higher hydraulic pressure prevails.
The vanes 152 designed to operate within the retarding hydraulic chamber 153 and the advancing hydraulic chamber 154 (i.e., operation range of the vane) and shifted in the angular position or phase are operatively coupled to the cam shafts 15C and 16C for driving the intake or exhaust valves of the engine cylinders.
Although not shown in the drawings, the actuator 16 on the exhaust side is provided with a spring for resiliently urging the vane 152 so that it can assume the advanced position against the reaction force of the cam shaft 16C.
The actuators 15 and 16 are driven under the hydraulic pressure of a lubricant oil of the engine 1 supplied through the oil control valves 19 and 20. For controlling the cam angle phases of the actuators 15 and 16 in such manner as illustrated in FIGS. 14 to 16, the amount of oil (i.e., hydraulic pressure) fed to the actuators 15 and 16 is controlled.
By way of example, regulation of the cam angle phase to the most retarded position, as illustrated in FIG. 14, can be realized by feeding oil into the retarding hydraulic chamber 153. On the contrary, regulation of the cam angle phase to the most advanced position, as illustrated in FIG. 16, can be effectuated by feeding oil into the advancing hydraulic chamber 153.
The oil control valves 19 and 20 are in charge of selecting either the retarding hydraulic chamber 153 or the advancing hydraulic chamber 154 for the oil supply. FIG. 17, 18 and 19 show in side-elevational sectional views the internal structures of the oil control valves 19 and 20 which are implemented substantially identically.
Referring to FIGS. 17 to 19, each of the oil control valves 19 and 20 is comprised of a cylindrical housing 191, a spool 192 slideably disposed within the housing 191, a solenoid coil 193 for driving continuously the spool 192 and a spring 194 for resiliently urging the spool 192 in the restoring direction.
The housing 191 is provided with an orifice 195 which is hydraulically communicated to a pump (not shown), orifices 196 and 197 hydraulically connected to the actuator 15 or 16, and drain orifices 198 and 199 fluidly communicated to an oil pan.
The orifice 196 can be communicated to the retarding hydraulic chamber 153 of the actuator 15 or the advancing hydraulic chamber 154 of the actuator 16. On the other hand, the orifice 197 can be communicated to the advancing hydraulic chamber 154 of the actuator 15 or the retarding hydraulic chamber 153 of the actuator 16.
The orifices 196 and 197 are selectively put into communication with the oil feeding orifice 195 in dependence on the axial position of the spool 192 (i.e., the position of the spool in the longitudinal direction thereof). In the state shown in FIG. 17, the orifice 195 is shown as having been placed in communication with the orifice 196, while in FIG. 19, the orifice 195 is shown as being communicated to the orifice 197.
Similarly, the drain orifices 198 and 199 are selectively put into communication with the orifice 197 or 196 in dependence on the axial position of the spool 192. In the state shown in FIG. 17, the orifice 197 is shown as being communicated with the orifice 198, while in FIG. 19, the orifice 196 is being communicated to the orifice 199.
The oil feed port formed in the locking recess 157 is so arranged as to be supplied with oil when the oil control valves 19 and 20 are in the electrically driven state (see FIG. 19). More specifically, when the hydraulic pressure applied to the locking recess 157 exceeds the spring force of the spring 156, the lock pin 155 is pushed out from the locking recess 157, whereby the locked state is cleared.
FIG. 17 shows the state in which the electric current flowing through the solenoid or coil 193 is at a minimum value and thus the spring 194 is stretched or relaxed to a maximum extent.
Assuming that the oil control valve shown in FIG. 17 serves as the oil control valve 19 of the intake side, the hydraulic medium or oil supplied from the pump via the orifice 195 flows into the retarding hydraulic chamber 153 of the actuator 15, as a result of which the actuators 15 is shifted to the state illustrated in FIG. 14.
Consequently, the oil resident in the advancing hydraulic chamber 154 of the actuator 15 is forced to flow out through the orifice 197 to be finally discharged to the oil pan by way of the orifice 198.
On the other hand, assuming that the oil control valve shown in FIG. 17 serves as the oil control valve 20 on the exhaust side, the situation is reversed. Namely, the hydraulic medium or oil supplied from the pump via the orifice 196 flows into the advancing hydraulic chamber 154 of the actuator 16, as a result of which the actuators 16 is ultimately set to the state illustrated in FIG. 16.
In that case, the oil contained in the retarding hydraulic chamber 153 of the actuator 16 is forcibly discharged to the oil pan by way of the orifices 197 and 198.
By virtue of the hydraulic circuit arrangement described above by reference to FIG. 17, valve overlap can be suppressed to a minimum even upon occurrence of failure such as shutdown of electric current supply to the oil control valves 19 and 20 disposed at the intake side and the exhaust side, respectively, due to wire breakage or the like. This feature is advantageous from the viewpoint of ensuring high withstandability against the engine stall.
In FIG. 19, the state is illustrated in which where the current flowing through the coil 193 is of a maximum value and thus the spring 194 is compressed to the minimum length.
Assuming, by way of example, that the oil control valve shown in FIG. 19 serves as the oil control valve 19 installed on the intake side, the oil fed from the pump is caused to flow into the advancing hydraulic chamber 154 of the actuator 15 via the orifice 197, whereas the oil in the retarding hydraulic chamber 153 of the actuator 15 is discharged via the orifices 196 and 199.
On the other hand, in the case where the oil control valve shown in FIG. 19 serves as the oil control valve 20 on the exhaust side, the oil fed from the pump is forced to flow into the retarding hydraulic chamber 153 of the actuator 16 via the orifice 197, while the oil in the advancing hydraulic chamber 154 of the actuator 16 is discharged via the orifices 196 and 199.
FIG. 18 shows the state corresponding to the valve timing control end position or lock-up position (mid position). In this state, the vanes 152 of the actuators 15 and 16 are at desired positions, respectively, (see the state illustrated in FIG. 15).
In the state illustrated in FIG. 18, the orifice 195 provided at the oil supply side is not directly communicated to the orifice 196 or 197 disposed on the actuator side. However, due to oil leakage, oil is supplied to the oil feed port of the locking recess 157 (see FIG. 15).
Accordingly, even when the vane 152 is at the lock-up position, there may arise such situation in which the hydraulic pressure applied to the oil feed port under the oil leakage overcomes the spring force of the spring 156 (i.e., exceeds the predetermined unlocking hydraulic pressure value). In that case, the lock pin 155 is caused to disengage from the locking recess 157, allowing the vane 152 to move or operate within the housing 151.
At this juncture, it should be mentioned that the predetermined unlocking hydraulic pressure mentioned above may be set at a necessary minimum value.
Furthermore, the positions (phases) of the vanes 152 of the actuators 15 and 16 which play the role for determining the valve timing can appropriately be controlled by detecting the vane positions by means of the cam angle sensors 17 and 18.
The cam angle sensors 17 and 18 are mounted at the positions which enable these sensors to detect the relative position between the crank shaft on one hand and the cam shafts 15C and 16C on the other hand.
Referring to FIG. 13, the phase difference relative to the output signal of the crank angle sensor at the position where the valve timing is most advanced (see the broken line curve shown in FIG. 12) is indicated by A, whereas the phase difference relative to the output signal of the crank angle sensor at the position where the valve timing is most retarded (see the single-dotted line curve shown in FIG. 12) is indicated by B.
The ECU 21 is designed or programmed to perform the feedback control so that the phase difference A or B as detected coincides with the desired value, whereby the valve timing control is carried out at given positions.
More specifically, it is assumed, by way of example only, that on the intake side, the detected position of the cam angle sensor 17 relative to the detection timing of the crank angle sensor 14 is retarded with reference to the desired position arithmetically determined by the ECU 21. In that case, the detected position (detection timing) of the cam angle sensor 17 has to be to advanced the desired position. To this end, the amount of the electric current flowing through the coil 193 of the oil control valve 19 is regulated in dependence on the difference between the detected position and the desired position, to thereby control correspondingly the spool 192.
Further, in the case where the difference between the desired position and the detected position is large, the amount of electric current supplied to the coil 193 of oil control valve 19 is increased in order to allow the desired position to be attained speedily.
As a result of this, the aperture of the orifice 197 opened into the advancing hydraulic chamber 154 of the actuator 15 is increased, which results in increasing of the amount of oil fed to the advancing hydraulic chamber 154.
Subsequently, as the detected position approaches to the desired position, the current supply to the coil 193 of the oil control valve 19 is decreased so that the position of the spool 192 of the oil control valve 19 becomes closer to the state illustrated in FIG. 18.
At the time point when coincidence is found between the detected position and the desired position, the electric current supply to the coil 193 is so controlled that the oil flow path leading to the retarding hydraulic chamber 153 and the advancing hydraulic chamber 154 of the actuator 15 is intercepted, as can be seen in FIG. 18.
Incidentally, the desired position in the ordinary engine operation mode (e.g. running state succeeding to the warm-up operation) can be so set or established that optimal valve timing can be realized in accordance with the engine operation state by previously storing, for example, two-dimensional map data values obtained experimentarily in correspondence to the operating states (e.g. engine rotation speeds (rpm) and engine loads), respectively, in a read-only memory or ROM incorporated in the ECU 21.
On the other hand, in the engine starting operation mode, the rotation speed of the oil pump which is driven by the engine 1 is not sufficiently high. Consequently, the volume of the oil fed to the actuator 15 is also insufficient. Thus, the control of the valve timing to the advanced position by controlling the hydraulic pressure as described previously is rendered practically impossible.
Such being the circumstances, jolting or fluttering of the vane 152 due to shortage of the hydraulic pressure has to be prevented by engaging the lock pin 155 with the locking recess 157, as illustrated in FIG. 15.
In that case, if the intake valve is actuated excessively retardingly (i.e., if the valve timing is overretarded), the actual compression ratio becomes lowered while excessive advancing of actuation of the intake valve (i.e., overadvancing of the valve timing) will result in increasing of the time period during which the intake valve and the exhaust valve overlap with each other. In other words, overretarded or overadvanced actuation of the intake valve results in increasing of the pumping loss.
Certainly, the overretarding or overadvancing actuation control of the intake valve can profitably be adopted for increasing the rotation speed in the engine starting operation (e.g. upon cranking) and triggering the initial explosion. However, because the combustion is essentially inadequate, complete combustion or explosion is difficult to realize.
On the other hand, overretarding of actuation of the exhaust valve will result in increasing of the overlap period during which the intake valve and the exhaust valve overlap with each other, similarly to the case where operation of the intake valve is advanced excessively. By contrast, overadvancing of the exhaust valve actuation will incur lowering of the actual expansion ratio, rendering it impossible to transmit the combustion energy sufficiently to the crank shaft.
As is apparent from the above, overretarding or overadvancing control of the valve timing in the engine starting operation or immediately thereafter may unwantedly incur degradation of the engine starting performance or the state incapable of starting the engine operation in the worst case.
Thus, for coping with the problems such as mentioned above in the engine starting operation, the vane 152 is fixedly set at the lock-up position (i.e., nearly mid position between the most retarded position and the most advanced position) by engaging the lock pin 155 into the locking recess 157, as shown in FIG. 15.
In that case, since the hydraulic pressure of the lubricating oil increases as the engine rotation speed (rpm) increases in succession to starting operation of the engine, the hydraulic pressure is fed to the actuators 15 and 16 because of the oil leakage described previously even in the state where the spool 192 is at the position shown in FIG. 18.
Such being the circumstances, when the hydraulic pressure applied to the locking recess 157 overcomes the spring force of the spring 156, the lock pin 155 is caused to disengage from the locking recess 157, allowing the vane 152 to move.
Thus, by controlling the oil control valves 19 and 20 after unlocking of the vanes, the hydraulic pressure fed to the retarding hydraulic chamber 153 and the advancing hydraulic chamber 154 can be regulated, whereby the valve timing retarding or advancing control can be carried out.
In that case, particularly in the high-speed rotation range of the engine 1, the valve timing is so controlled as to be retarded more when compared with the engine starting operation for the purpose of realizing the suction inertia effect as well as enhancement of the volumetric efficiency and hence the output performance of the engine.
As can be appreciated from the foregoing, in the engine starting operation, the lock pins 155 of the actuators 15 and 16 are locked at a nearly mid position between the most retarded position and the most advanced position with a view to enhancing the engine starting performance. On the other hand, once the engine operation has been started after releasing of the locking mechanism, the valve timing is so controlled as to be retarded especially in the high-speed rotation range of the engine.
However, in the conventional valve timing control system for the internal combustion engine, no consideration has been given to such technical matters as improvement of the exhaust gas quality and promotion of temperature rise of the catalyst.
The conventional valve timing control system for an internal combustion engine is configured as described above. In an engine starting operation, the valve timing control system engages with a substantially intermediate position between a most advanced position and a most retarded position by the locking mechanism of the actuator, thereby improving a starting performance of the internal combustion engine. After the engine operation has been started, when the locking mechanism is released, the valve timing control system improves an output performance of the internal combustion engine by controlling the valve timing toward a more retarded state than in the starting operation, in particular, in a high rotational range.
In addition, Japanese Patent Application Laid-open No. Hei 11-210424 describes that, after the lock pin is released, a control of valve timing executes a feedback control for making a detected advance angle amount coincide with a target advance angle amount.
On the intake side, if the detected advance angle amount is in a more retarded state than the target advance angle amount, the valve timing control system controls the OCVs 19 and 20 to supply oil to the advancing hydraulic chamber of the actuator in order to advance the valve timing. As a result, the OCVs are capable of successively controlling the spool 192 to be set to an arbitrary position by an amount of energizing current to the coil 193 as shown in FIG. 19, thereby successively controlling an amount of oil to be supplied from an oil pump to the actuators 15 and 16.
If the detected advance angle amount is in a more advanced state than the target advance angle amount, the valve timing control system controls the OCVs to supply oil to the retarding hydraulic chamber of the actuator as shown in FIG. 17 so that the valve timing is retarded. In addition, if the detected advance angle amount substantially coincides with the target advance angle amount, the valve timing control system controls both the advancing hydraulic chamber 154 and the retarding hydraulic chamber 153 to be set to positions for blocking a passage as shown in FIG. 18.
If the target advance angle amount is in a pin-lock-up position, the lock pin 155 is in the position of the locking recess 157, and most of the passages of the OCVs 19 and 20 are blocked. Thus, since a hydraulic pressure decreases by a large degree and a hydraulic pressure applied to the lock pin 155 also decreases, the lock pin 155 is locked in the locking recess 157 if a force caused by the hydraulic pressure applied on the lock pin 155 becomes smaller than a spring force.
Here, in the case in which an integral control is executed in order to make the detected advance angle amount coincide with the target advance angle amount, the detected advance angle amount is locked by the lock pin 155 if there is only a slight difference between the pin lock-up position and the target advance angle amount when the lock pin 155 is locked. Thus, the detected advance angle amount does not move despite the fact that an integrated value is increased or decreased, and the integrated value is increased or decreased to a limit of a control range. When the target advance angle amount changes and it is intended to make the detected advance angle amount follow the change, the detected advance angle amount may not be able to follow the target advance angle amount promptly because a control value diverges.
In addition, when passages to the actuators of the OCVs are secured and a hydraulic pressure to the lock pin 155 reaches a hydraulic pressure sufficient to release a lock before the integrated value reaches the limit of the control range, the pin lock is released and a control amount deviates largely due to a movement of the integrated value at this point. Thus, the detected advance angle amount may deviate largely from the target advance angle amount simultaneously with the release of the lock pin.
The present invention has been devised in order to solve the above and other drawbacks, and it is an object of the present invention to realize a valve timing control system of an internal combustion engine for preventing divergence of a control amount and an unexpected release of a lock pin in the case in which a target advance angle amount or a detected advance angle amount is controlled to be set substantially to a pin lock-up position, and at the same time preventing a detected advance angle amount from deviating from a target advance angle amount even in the case in which a lock pin is released with a control amount dispersing, thereby eliminating a decrease of an engine performance to prevent decrease of drivability, a mileage, a gas exhausting performance or the like.
In view of the above and other objects which will become apparent as the description proceeds, there is provided according to an aspect of the present invention a valve timing control system for an internal combustion engine, which system includes sensor means for detecting engine operation states of an internal combustion engine, intake or exhaust cam shafts for driving intake or exhaust valves, respectively, of the internal combustion engine in synchronism with a rotation of a crank shaft of the internal combustion engine, at least one actuator operatively connected to at least one of cam shafts for driving the intake or exhaust valves, respectively, a hydraulic pressure supply unit for feeding a hydraulic pressure to drive the actuator; and control means for controlling the hydraulic pressure fed from the hydraulic pressure supply unit to the actuator in dependence on the operating states of the internal combustion engine while changing a relative phase of the cam shaft relative to the crank shaft, wherein the actuator includes a retarding hydraulic chamber and an advancing hydraulic chamber for setting an adjustable range of the relative phase, a locking mechanism for setting the relative phase to a lock-up position within the adjustable range, and an unlocking mechanism for releasing the locking mechanism in response to a predetermined level of hydraulic pressure fed from the hydraulic pressure supply unit, and wherein, when driving the locking mechanism to control the relative phase to be within a predetermined range of the lock-up position, the control means reduces a limit of a control range.
Further, the control means detects a detected advance angle amount that is a phase difference between phases of the crank shaft and the cam shaft, and calculates a target advance angle amount that is a valve timing suited for an operating state of the internal combustion engine to make a limit of control range of an integrated value to be smaller than in the case in which the detected advance angle amount is not in the lock-up position if the detected advance angle amount is subject to an integral control to be substantially coincident with the target advance angle amount.
Furthermore, the control means initializes an integrated value if the target advance angle amount or the detected advance angle amount is changed to the outside of a predetermined range from being within a predetermined range of a lock-up position in the locking mechanism.
Still further, the control means executes the initialization of the integrated value only when the integrated value reaches the limit of control range.
Yet still further, the control means does not make the limit of control range smaller if a period when the target advance angle amount or the detected advance angle amount is within a predetermined range of a lock-up position in the locking mechanism is within a predetermined period.
Furthermore, the period when the target advance angle amount or the detected advance angle amount is within a predetermined range of a lock-up position in the locking mechanism is a period until the integrated value reaches the limit of control range.
In addition, the control means stops the integral control if the target advance angle amount or the detected advance angle amount is within a predetermined range of a lock-up position in the locking mechanism.
Finally, the control means executes the controls only when the engine operation states of the internal combustion engine is in a predetermined operating state.